Numerous problems have arisen for both constant air volume and variable air volume systems due to the efforts to reduce the operating cost, reduce the capital cost of installations and reduce the space requirements for the air conditioning systems. While some of these problems have been successfully resolved, others have been solved by means which have largely nullified the original design objectives and frequently degraded performance to an unacceptable level, especially in offsetting latent heat loads.
In addressing the imposed loads the design of costeffective quality air conditioning systems requires, inter alia, consideration of
(i) Coolant flow rate and air flow rate, PA1 (ii) Dehumidifier size, PA1 (iii) Secondary to primary surface area ratio, PA1 (iv) Performance at all potential operating conditions, PA1 (v) System noise levels. PA1 (a) Range of load PA1 (b) Segments of that range PA1 (c) Stages of effective dehumidification PA1 (b) Coolant Flow Rate and Variable Air Volume (VAV) PA1 Systems
In this specification, consideration is given to:
These requirements are referred to hereunder in detail:
(i) Coolant Flow Rate and Air Flow Rate PA0 (a) Conventional Coolant Flow Rate for Constant Air PA0 Volume (CAV) Systems PA0 (ii) Dehumidifier Size PA0 (iii) Secondary to Primary Surface Area Ratio, (Fin PA0 Density). PA0 (iv) Performance
The flow rate of coolant influences part load performance in all environments. The higher the coolant velocity within the tubes of the dehumidifier, all other parameters being held constant, the steeper is the coil condition curve on a psychrometric chart; that is, the greater is the ratio of latent cooling (moisture removal) to sensible cooling.
Conventionally, whether the air conditioning system is a constant air volume (CAV) system or a variable air volume (VAV) system, it is common practice to effect control by reducing the volume flow rate of coolant through the tubes of the dehumidifier coil as the cooling requirement reduces. This reduces the cooling capacity of the coil but also reduces the ratio of the latent to sensible cooling by reducing the coolant-side heat transfer coefficient, itself a function of coolant flow velocity, so raising the coil surface temperature, hereinafter referred to as the interface temperature.
During part load weather conditions the transmission of sensible heat from the external environment to the treated zone reduces, or may actually become negative and so canoel part of the internal sensible heat load. However latent heat addition (from people, infiltration and other sources) which occurs simultaneously and in parallel with the sensible transfer, will usually remain the same or may increase It is quite common to have a part load condition wherein the ambient dry bulb temperature is lower and the dew point temperature is higher than at design peak conditions. Thus there is a decreased sensible heat load and an increased latent heat load. The dehumidifier must then operate at a new ratio of latent to sensible heat transfer and hence the slope of the coil condition curve is required to be steeper.
Whether the system be constant air volume (CAV) or a variable air volume (VAV) the velocity of the airstream entering the face of the dehumidifier coil, hereinafter referred to as the face velocity, influences performance. The lower the face velocity, all other parameters including mass flow of air being held constant, the lower is the air-side heat transfer coefficient, the lower is the coil surface (interface) temperature, the greater is the amount of moisture removal per unit mass flow of air, the greater is the ratio of latent to sensible cooling and hence the steeper and straighter is the coil condition curve.
In constant air volume systems the conventional face velocity does not vary with the load. A reduced load is offset by throttling the coolant flow to the dehumidifier. As a result of the decrease in heat transfer rate due to reduced coolant flow, which for a given coolant circuit arrangement and series connection of all coil portions is synonymous With reduced coolant velooity, the air temperature leaving the dehumidifier rises with throttling of the coolant flow. This can only be a satisfactory means of accommodating reduced loads if the zone latent heat loads are small and the ambient air at part load is dry.
Otherwise, the reduced coolant flow allows the interface temperature to rise as a result of the decrease in coolant-side heat transfer coefficient, which in turn reduces the rate of moisture removal from the air and causes the slope of the coil condition curve to decrease such that the ratio of latent to sensible heat transfer decreases below that for full load. To satisfy the ratio required for a particular part load condition the dew point of the air entering the dehumidifier must increase to provide a sufficient difference from the interface temperature to cause condensation to occur at the required rate. This in turn requires that the humidity ratio in the conditioned space must rise. Often the level to which it rises is unacceptable to the occupants of the space. As the throttling of the coolant proceeds, the humidity ratio of the air leaving the dehumidifier rises progressively. However, it has already been established that during part load for a given entry condition a steeper coil condition curve is required to accommodate the increased ratio of the latent to the sensible heat load. It is evident also that in climates having high humid peak load conditions steep coil condition curves are required.
In basic VAV systems the leaving supply air temperature is generally kept constant and the flow rate of air is reduced as the sensible load reduces. As for the constant air volume system, the coolant flow is also throttled and again this tends to decrease the slope of the coil condition curve for a given air entering condition since the coolant-side heat transfer coefficient is reduced. However this effect is partially offset by the reduction in the air flow rate, which reduces the air side heat transfer coefficient and, as discussed above and illustrated in FIG. 7a, also reduces the interface temperature of the air and the interface temperature over a larger proportion of the coil, resulting in an improved driving force for dehumidification. The combined result of these two opposing influences is that throttling of the coolant flow rate at part load causes the slope of the coil condition curve for a given air entering condition in a VAV system to be steeper than that in a CAV system but less steep than that achieved by the present invention. Reducing the coolant temperature rise by careful choice of coolant flow and flow circuiting, and/or lowering the coolant supply temperature, are additional means by which the steepness of the coil condition curve may be controlled.
The mismatch which exists between the size of the dehumidifier coil selected for full load design conditions and the actual load to be offset at part load conditions constitutes one major difficulty which is overcome by this invention.
It is not uncommon for an air conditioning system to be required to satisfy a part load sensible condition which is 40% or 30% of the full design sensible load. Existing practice appears not to appreciate the consequences which result when a dehumidifier, which is properly sized for a peak design load, is required to perform at part load conditions. It is rare for part load performance to be specified by consulting engineers. At low load conditions the coolant flow rate through a given coil, which for such conditions is disproportionately large in relation to the magnitude of the load, drops to a trickle. Inevitably, the heat transfer coefficient inside the tubes reduces to a small value and the coil surface temperature increases.
The reduction in the coolant side heat transfer coefficient occurs both with liquid flow coolants such as chilled water or ethylene glycol, and with liquid and vapour flow coolants such as refrigerant R12 or R22. In the latter case a number of flow patterns occur depending on the mass fraction of liquid, the fluid properties of each phase and the flow rate. A good understanding of the effect of low mass velocities of refrigerants on the heat transfer coefficient is presented in FIG. 9 ASHRAE Handbook 1985, Fundamentals, published by the American Society of Heating Refrigerating and Air-Conditioning Engineers Inc., Atlanta, Georgia, U.S.A., on p 4.7. It is there clearly demonstrated that a drop in the mass flow rate of the refrigerant to 40% of the indicated peak mass flow rate is associated with a drop of up to 34% in the heat transfer coefficient.
For a large proportion of the coil the surface temperature may become greater than the dew point temperature of the air to be treated, with a consequent loss of dehumidification. For this second reason the slope of the coil condition curve of a conventional air conditioning system at part loads becomes shallow just when it is required to become steep, despite the steepening effect of a drop in face velocity of air passing through the coil.
The lower the temperature of the wetted outside surfaces of the coil the greater will be the condensation of water vapour on those surfaces. Fins, or secondary surfaces, have a higher surface temperature than do the tubes, or primary surfaces. As fin density increases, the average fin temperature also increases. By having a large proportion of primary surface area, the dehumidification per unit of surface area will be large; but if taken too far this consideration would lead to coils with many rows of depth which do not make efficient use of the material of which they are made. Thus there is an optimum ratio of secondary to primary surface which gives the best use of material in achieving the required degree of dehumidification for a given application. Seeking to reduce coil depth by using very high fin density is poor practice if dehumidification is required. While it may result in a small reduction in size and therefore first cost of the dehumidifier, there is firm evidence that it inhibits dehumidification and hence compromises part load performance. The slope of the coil condition curve will decrease, performance will be impaired and fan power requirements will be increased because of the higher resistance offered to the air flow by the high fin density.
The variable air volume (VAV) system is frequently employed in air conditioning design, especially when energy consumption and space requirements are considered. However the system has often been widely criticized by building occupants because under part load conditions performance does not satisfy expectations. One article by Tamblyn in the September 1983 ASHRAE Journal, with reference to new VAV systems, lists complaints of "... stale air and lack of air motion ..." and reports that "Owners are fighting back in energy consuming ways by raising outside air ratios, operating fans longer and setting minimum airflows which demand the use of the same reheat that was formerly eliminated".
Reference can also be made to the August 1987 ASHRAE Journal, page 22, wherein the problems of VAV systems are discussed in detail by a distinguished forum. The problems are listed as uneven temperatures, lack of temperature and humidity controls, lack of air motion, lack of fresh air, and excessive energy consumption. Even reheating is recommended in that article as a realistic solution to the problems. Further, it has been suggested therein that only interior zones should be serviced by VAV systems.
A typical VAV system which is particularly advantageous in conserving both space and energy is that in a high rise office block which employs air handling units on each floor. The need for large shaft spaces and long duct runs is eliminated since each air handling unit is located on the floor it serves. It is conventional to utilize the ceiling space as a large return air plenum. If such a building is located in a city, such as Melbourne, Australia, or Dallas, Texas, the system will be designed to operate with a high outside air dry bulb temperature, say 95.degree. F. (35.degree. C.) and a low humidity during summer peak design conditions. During part load days and marginal weather conditions when the ambient dry bulb temperature is lower, there are numerous periods during which the humidity ratio is considerably above the summer peak conditions. A typical minimum fresh air intake is the equivalent of 15% of the total peak design airflow rate. Since the minimum fresh air intake for meeting ventilation requirements is a fixed quantity, at 60% part load the requirement for outside air is (15/0.6)%, i.e. 25%, and at 30% part load the requirement is for 50% outside air. Thus the dehumidifier is burdened on humid part load days not only with an outside air humidity ratio condition which is higher than that at peak loads, but also with a higher percentage of outside air. Frequently this demand is beyond the capability of the conventional VAV system which largely accounts for the many complaints that the atmosphere is "humid" or "stuffy".